There is known an exhaust hood which is positioned immediately after the last-stage blades of a turbine downstream of the working fluid.
The walls of the exhaust casing define a passage which is formed as an axial-radial diffuser at the inlet, and as a channel of rectangular cross section at the outlet.
Positioned at the passage inlet is a guide member which is provided to ensure without-separation flow of a working fluid during its turning or spreading in the axial-radial diffuser. The guide member is formed as a body of revolution, whose axis is aligned with that of the turbine, with one end thereof disposed in proximity with the last-stage blades of the turbine.
Positioned in the passage downstream from the guide member is a deflector which is also formed as a body of revolution, whose axis is in parallel with that of the turbine. The deflector is provided to permit aerodynamic force to act on the flow by altering the shape and flow area of the annular channels formed by the guide member and deflector, and by the deflector with the inside wall of the exhaust hood passage. This is achieved by that the deflector is displaced by means of a driving mechanism either in longitudinal and/or transverse directions relative to the turbine axis.
The exhaust hood construction described above fails to provide a required efficiency in the exhaust hood at under-load performance of the turbine, and does not ensure sufficient operating reliability of the turbine in the event the working fluid rates being less than 1/3 of the nominal value.
It is known to those skilled in the art that "the rated load" of the turbine (and "the nominal working-fluid rate" which corresponds to the latter) is the load which assures maximum efficiency and operating reliability of the turbine.
With a decrease in the turbine load, followed by a decrease in the rates of a working fluid passing through the last stage of the turbine at a constant rotational speed of the turbine last-stage blades, the flow of the working fluid enters the exhaust hood as a rotating body. The circumferential velocity component is in general comparable with the axial velocity component, but in certain instances it may be considerably greater. The centrifugal forces in the rotational flow make for an increase in the radial velocity component, and for separation of the flow from the inside wall of the exhaust hood, whereby the efficiency and operating reliability of the turbine are impaired.
Therefore, by changing the flow areas of the annular channels at the inlet portion of the exhaust hood, it becomes possible to limit any further increase in the radial velocity component directly in the last stage of the turbine and after this stage. In this way it becomes feasible to obtain axi-symmetrical without-separation flow of the working fluid with the rates thereof ranging from nominal to approximately 1/3 of the nominal. However, with the working fluid rates being less than 1/3 of the nominal, even substantial changes in the flow area of the above-mentioned channels will not prevent the flow from separation from the inside wall of the axial-radial diffuser of the passage and directly in the last-stage blades, which impairs the operating reliability of the turbine.
Moreover, in the prior-art exhaust hood construction, with lower rated loads of the turbine, the flow of working fluid passes beyond the boundary of the axial-radial diffuser as a body rotating over the entire area of the diffuser, which leads to substantial mechanical losses of the flow energy required for the vortex formation in the main portion of the exhaust hood.
It has been experimentally found that the greater part of the working-fluid particles travel along the trajectories known to be non-optimal.
Thus, the prior-art exhaust hood not only fails to restore static pressure of the working fluid therein, but permits a drop in this pressure to take place from the exhaust inlet to its outlet, which reduces the efficiency of the turbine as a whole.